Engine controller

ABSTRACT

A controller for an engine estimates a temperature of the exhaust gas and controls the engine according to the estimated exhaust temperature. The controller changes the air-fuel ratio to a stoichiometric air-fuel ratio or leaner. The controller calculates the progress of combustion on the basis of signals of sensors, and estimates an exhaust temperature. In the case where the air-fuel ratio is the stoichiometric air-fuel ratio, the controller estimates the exhaust temperature on the basis of the progress of the combustion, the engine temperature, and a first relationship that is at least defined between the progress of the combustion and the exhaust temperature, . In the case where the air-fuel ratio is lean, the controller estimates the exhaust temperature on the basis of the progress of the combustion, the engine temperature, and a second relationship that differs from the first relationship.

TECHNICAL FIELD

A technique disclosed herein relates to an engine controller.

BACKGROUND ART

In Patent document 1, a technique of estimating a temperature of exhaustgas in a spark-ignition engine is described. More specifically, in thistechnique, an electronic control unit uses an exhaust temperature map,in which a relationship between each of an engine load and an enginespeed and the temperature of the exhaust gas is defined, so as toestimate the temperature of the exhaust gas on the basis of an operationstate of the engine. The temperature of the exhaust gas is affected byan ignition timing. Accordingly, in Patent document 1, it is describedthat the estimated temperature of the exhaust gas is corrected by acorrection coefficient that is defined according to a difference betweencombustion gravity center at the time when the ignition timing is amaximum brake torque (MBT) and the combustion gravity center of actualcombustion. The correction coefficient is set to be increased in theform of a quadratic curve as the difference in the combustion gravitycenter is increased.

In addition, in Patent document 1, it is described that the temperatureof the exhaust gas is changed by an air-fuel ratio of an air-fuelmixture. More specifically, in Patent document 1, it is described that afunction between a reference exhaust temperature and an air-fuel ratiois set and that a second correction coefficient is calculated accordingto this function.

PRIOR ART DOCUMENTS Patent Documents

Patent document 1: JP 2017-223138A

SUMMARY OF THE INVENTION Problem to be Solved by the Invention

The temperature of the exhaust gas is determined from a quantity ofheat, which is acquired by subtracting a quantity of heat used fordriving of the engine (that is, illustrated work) and a quantity of heatreleased to the engine (that is, cooling loss) from a quantity of heatgenerated by combustion in a cylinder.

Here, it is understood from the research by the present inventors andthe like that the cooling loss was changed with a change in the air-fuelratio of the air-fuel mixture. When it is desired to accurately estimatethe temperature of the exhaust gas, unlike in Patent document 1, notonly the air-fuel ratio of the air-fuel mixture but also the coolingloss has to be considered.

A technique disclosed herein estimates a temperature of exhaust gasaccurately in an engine that changes an air-fuel ratio of air-fuelmixture.

Means for Solving the Problem

A technique disclosed herein relates to an engine controller for anengine. This controller includes: an exhaust passage which is connectedto the engine and through which exhaust gas is discharged from inside ofa cylinder of the engine; a sensor which outputs a signal correspondingto a combustion state in the cylinder; and a control unit to which thesensor is connected, which estimates a temperature of the exhaust gas onthe basis of the signal of the sensor, and which controls the engineaccording to the estimated temperature of the exhaust gas.

The control unit changes an air-fuel ratio of an air-fuel mixture in thecylinder to a stoichiometric air-fuel ratio or a leaner air-fuel ratiothan the stoichiometric air-fuel ratio according to an operation stateof the engine.

The control unit includes a processor configured to execute acalculation module that calculates progress of combustion as a crankangle at the time when the combustion in the cylinder is progressed to aparticular extent on the basis of the signal of the sensor; and anestimation module that estimates the temperature of the exhaust gas onthe basis of the progress of the combustion calculated by thecalculation module, the air-fuel ratio of the air-fuel mixture, and atemperature of the engine.

In a case where the air-fuel ratio of the air-fuel mixture is thestoichiometric air-fuel ratio, the estimation module estimates thetemperature of the exhaust gas on the basis of the progress of thecombustion, the temperature of the engine, and a first relationship thatis at least defined between the progress of the combustion and thetemperature of the exhaust gas, and in a case where the air-fuel ratioof the air-fuel mixture is leaner than the stoichiometric air-fuelratio, the estimation module estimates the temperature of the exhaustgas on the basis of the progress of the combustion, the temperature ofthe engine, and a second relationship that differs from the firstrelationship and is at least defined between the progress of thecombustion and the temperature of the exhaust gas.

The engine controller with this configuration estimates the temperatureof the exhaust gas on the basis of the progress of the combustion. Morespecifically, the calculation module of the control unit calculates theprogress of the combustion that corresponds to the crank angle at thetime when the combustion in the cylinder is progressed to the particularextent on the basis of the signal of the sensor. The sensor outputs thesignal corresponding to the combustion state in the cylinder. The sensormay be an in-cylinder pressure sensor that outputs a signalcorresponding to an in-cylinder pressure. The control unit canaccurately comprehend the combustion state in the cylinder on the basisof the signal of the in-cylinder pressure sensor.

The progress of the combustion is the crank angle at the time when thecombustion is progressed to the particular extent. Thus, the progress ofthe combustion can be used as a parameter representing the combustionstate. The crank angle at which a mass combustion rate acquires aparticular value, for example, the crank angle at which the masscombustion rate is 50% (that is, mass fraction burned 50: mfb50) may beused as the progress of the combustion. The term “mfb50” means the crankangle at which 50% of a total injection amount of the fuel is burned.However, the progress of the combustion is not limited to mfb50. As longas there is a correlation between the progress of the combustion and thetemperature of the exhaust gas, any value such as mfb10 or mfb90 can beused as the progress of the combustion.

The control unit changes the air-fuel ratio of the air-fuel mixtureaccording to the operation state of the engine. More specifically, thecontrol unit changes the air-fuel ratio of the air-fuel mixture to thestoichiometric air-fuel ratio or the leaner air-fuel ratio than thestoichiometric air-fuel ratio.

The estimation module of the control unit estimates the temperature ofthe exhaust gas on the basis of the progress of the combustion, which iscalculated by the calculation module, the air-fuel ratio of the air-fuelmixture, and the temperature of the engine. The temperature of theengine may be a temperature of an engine coolant, for example. In thecase where the air-fuel ratio of the air-fuel mixture is lean, thermalefficiency of the engine is relatively high, which reduces an amount ofthe fuel supply to the cylinder. As a result, the temperature inside thecylinder during the combustion is lower than that in the case where theair-fuel ratio of the air-fuel mixture is the stoichiometric air-fuelratio. In addition, cooling loss is less than that in the case where theair-fuel ratio of the air-fuel mixture is the stoichiometric air-fuelratio. That is, when the air-fuel ratio of the air-fuel mixture ischanged, the amount of the cooling loss is also changed. Thus, thetemperature of the exhaust gas is also changed by the change in theamount of the cooling loss. A relationship between the progress of thecombustion and the temperature of the exhaust gas is changed with achange in the air-fuel ratio of the air-fuel mixture. In addition, whenthe temperature of the engine is changed, the amount of the cooling lossis changed. Thus, the temperature of the exhaust gas is also changed.

The estimation module estimates the temperature of the exhaust gas onthe basis of the first relationship, which is at least defined betweenthe progress of the combustion and the temperature of the exhaust gas,the progress of the combustion, and the temperature of the engine in thecase where the air-fuel ratio of the air-fuel mixture is thestoichiometric air-fuel ratio. The estimation module estimates thetemperature of the exhaust gas on the basis of the second relationship,which is at least defined between the progress of the combustion and thetemperature of the exhaust gas, the progress of the combustion, and thetemperature of the engine in the case where the air-fuel ratio of theair-fuel mixture is leaner than the stoichiometric air-fuel ratio. Theestimation module switches between the first relationship and the secondrelationship according to the case where the air-fuel ratio of theair-fuel mixture is the stoichiometric air-fuel ratio or leaner than thestoichiometric air-fuel ratio. In this way, the estimation module canaccurately estimate the temperature of the exhaust gas in considerationof the temperature of the engine both in the case where the air-fuelratio of the air-fuel mixture is the stoichiometric air-fuel ratio andin the case where the air-fuel ratio of the air-fuel mixture is leanerthan the stoichiometric air-fuel ratio.

The estimation module may estimate the temperature of the exhaust gasfrom the progress of the combustion on the basis of the relationshipbetween the progress of the combustion and the temperature of theexhaust gas, and may correct the estimated temperature of the exhaustgas according to the temperature of the engine. The estimation modulemay change the correction amount of the temperature of the exhaust gasaccording to the air-fuel ratio of the air-fuel mixture.

In this way, the estimation module can change the correction amountaccording to the cooling loss, which is changed by the change in theair-fuel ratio of the air-fuel mixture. The estimation module canaccurately estimate the temperature of the exhaust gas on the basis ofthe progress of the combustion, the air-fuel ratio of the air-fuelmixture, and the temperature of the engine.

In the case where the air-fuel ratio of the air-fuel mixture is thestoichiometric air-fuel ratio, the estimation module may increase thecorrection amount of the temperature of the exhaust gas to be largerthan in the case where the air-fuel ratio of the air-fuel mixture isleaner than the stoichiometric air-fuel ratio.

As described above, in the case where the air-fuel ratio of the air-fuelmixture is lean, the amount of the cooling loss is relatively small. Onthe contrary, in the case where the air-fuel ratio of the air-fuelmixture is the stoichiometric air-fuel ratio, the amount of the coolingloss is relatively large.

Accordingly, in the case where the air-fuel ratio of the air-fuelmixture is the stoichiometric air-fuel ratio, the estimation moduleincreases the correction amount of the temperature of the exhaust gas.In the case where the air-fuel ratio of the air-fuel mixture is lean,the correction amount is reduced. In this way, the estimation module canaccurately estimate the temperature of the exhaust gas in considerationof the relationship between the air-fuel ratio of the air-fuel mixtureand the cooling loss.

The estimation module may make a correction to reduce the estimatedtemperature of the exhaust gas as the temperature of the engine isreduced in a case where the temperature of the engine is equal to orlower than a specified temperature, and may make a correction toincrease the estimated temperature of the exhaust gas as the temperatureof the engine is increased in a case where the temperature of the engineexceeds the specified temperature.

That is, each of the first relationship and the second relationship maybe set as a relationship between the progress of the combustion and thetemperature of the exhaust gas that is defined when the temperature ofthe engine is the specified temperature. The temperature of exhaust gas,which is estimated on the basis of the above relationship and theprogress of the combustion, corresponds to the temperature of theexhaust gas in the case where the temperature of the engine is thespecified temperature. In the case where the temperature of the engineis equal to or lower than the specified temperature, the amount of thecooling loss is increased as the temperature of the engine is reduced.Thus, the estimation module makes the correction to reduce the estimatedtemperature of the exhaust gas. In this way, the estimation module canaccurately estimate the temperature of the exhaust gas. Meanwhile, inthe case where the temperature of the engine exceeds the specifiedtemperature, the cooling loss is reduced as the temperature of theengine is increased. Thus, the estimation module makes the correction toincrease the estimated temperature of the exhaust gas. In this way, theestimation module can accurately estimate the temperature of the exhaustgas.

The control unit may switch between a first combustion mode, in whichthe air-fuel mixture in the cylinder is forcibly ignited according tothe operation state of the engine, so as to burn the air-fuel mixture byflame propagation, and a second combustion mode, in which the air-fuelmixture in the cylinder is forcibly ignited, so as to burn some of theair-fuel mixture by self-ignition.

In the second combustion mode, the control unit may change the air-fuelratio of the air-fuel mixture according to the operation state of theengine.

The present applicant proposes spark controlled compression ignition(SPCCI) combustion in which spark ignition (SI) combustion andcompression ignition (CI) combustion are combined. The SI combustion iscombustion that is initiated at the time when the air-fuel mixture inthe cylinder is forcibly ignited and that is associated with flamepropagation. The CI combustion is combustion that is initiated at thetime when the air-fuel mixture in the cylinder is self-ignited. TheSPCCI combustion is in a mode in which, when the air-fuel mixture in thecylinder is forcibly ignited to initiate the combustion by the flamepropagation, unburned air-fuel mixture in the cylinder is burned by theself-ignition due to a pressure increase caused by heat generation andthe flame propagation in the SI combustion.

The first combustion mode corresponds to a mode in which the SIcombustion is performed. The second combustion mode corresponds to amode in which the SPCCI combustion is performed. In the SI combustion,in the case where the air-fuel ratio of the air-fuel mixture is leanerthan the stoichiometric air-fuel ratio, combustion stability is possiblydegraded. Meanwhile, in the SPCCI combustion, some of the air-fuelmixture is burned by the self-ignition. Thus, even in the case where theair-fuel ratio of the air-fuel mixture is leaner than the stoichiometricair-fuel ratio, it is possible to stably burn the air-fuel mixture. Inthe second combustion mode, the air-fuel ratio of the air-fuel mixtureis changed according to the operation state of the engine. In this way,the engine can simultaneously secure combustion stability and improvethermal efficiency.

In a case where the temperature of the exhaust gas is higher than areference temperature, the control unit may increase an amount of thefuel supplied to the cylinder to be larger than in a case where thetemperature of the exhaust gas is equal to or lower than the referencetemperature.

When the amount of the fuel supplied to the cylinder is increased, dueto latent heat of the fuel, the amount of which is increased, thetemperature of the exhaust gas discharged from the cylinder is reduced.By reducing the temperature of the exhaust gas to be lower than thereference temperature, it is possible to secure reliability of acatalytic device provided in the exhaust passage of the engine, forexample. The performance of the catalytic device can be maintained in ahigh state. Thus, the engine can discharge purified gas.

In a case where the temperature of the exhaust gas is higher than areference temperature, the control unit may reduce a temperature of acoolant supplied to the engine to be lower than in a case where thetemperature of the exhaust gas is equal to or lower than the referencetemperature.

When the temperature of the coolant supplied to the engine is reduced,the amount of the cooling loss is increased. As a result, thetemperature of the exhaust gas discharged from the cylinder is reduced.Thus, similar to the above, it is possible to secure the reliability ofthe catalytic device.

In the case where the temperature of the exhaust gas is higher than areference temperature, the control unit may increase a flow rate of thecoolant supplied to the engine to be larger than in a case where thetemperature of the exhaust gas is equal to or lower than the referencetemperature.

When the flow rate of the coolant supplied to the engine is increased,the amount of the cooling loss is increased. As a result, thetemperature of the exhaust gas discharged from the cylinder is reduced.Thus, similar to the above, it is possible to secure the reliability ofthe catalytic device.

Advantage of the Invention

As it has been described so far, the engine controller can accuratelyestimate the temperature of the exhaust gas in the engine that changesthe air-fuel ratio of the air-fuel mixture.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a view exemplifying a configuration of an engine.

FIG. 2 is a block diagram exemplifying a configuration of an enginecontroller.

FIG. 3 include graphs, in which an upper graph is a graph exemplifying awaveform of SPCCI combustion and a lower graph is a graph exemplifyingan output signal of a crank angle sensor.

FIG. 4 is a graph exemplifying an operation map of the engine.

FIG. 5 include charts, each of which exemplifies fuel injection timing,ignition timing, and the combustion waveform in a respective operatingrange of the map in FIG. 4.

FIG. 6 is a block diagram exemplifying a functional block of an enginecontrol unit (ECU) that estimates a temperature of exhaust gas.

FIG. 7 is a flowchart exemplifying a procedure of estimating thetemperature of the exhaust gas that is executed by the ECU.

FIG. 8 is a graph exemplifying a model that represents a relationshipbetween progress of combustion and the temperature of the exhaust gas.

FIG. 9 include graphs, in which an upper graph is a graph exemplifyingmodels at the time when an engine speed is high and low in SIcombustion, and a lower graph is a graph exemplifying models at the timewhen the engine speed is high and low in the SPCCI combustion.

FIG. 10 is a graph exemplifying a relationship between the engine speedand the temperature of the exhaust gas.

FIG. 11 include graphs, in which an upper graph is a graph exemplifyingmodels at the time when an engine load is large and small in the SIcombustion, and a lower graph is a graph exemplifying models at the timeof when the engine load is large and small in the SPCCI combustion.

FIG. 12 is a graph exemplifying a relationship between a temperature ofa coolant and a correction amount of the temperature of the exhaust gas.

MODES FOR CARRYING OUT THE INVENTION

A detailed description will hereinafter be made on an embodiment of anengine controller with reference to the drawings. The followingdescription exemplifies the engine, an engine system, and the enginecontroller.

FIG. 1 is a view exemplifying a configuration of the engine system. FIG.2 is a block diagram exemplifying a configuration of the enginecontroller.

An engine 1 is a four-stroke engine that is operated by repeating anintake stroke, a compression stroke, a power stroke, and an exhauststroke in a combustion chamber 17. The engine 1 is mounted on afour-wheeled automobile. The automobile travels by operating the engine1. As the engine in this configuration example, the engine 1 usesgasoline as fuel. The fuel only has to be liquid fuel that at leastincludes gasoline. The fuel may be gasoline that contains bioethanol orthe like, for example.

(Engine Configuration)

The engine 1 includes a cylinder block 12 and a cylinder head 13 placedthereon. A plurality of cylinders 11 are formed in the cylinder block12. FIG. 1 illustrates only one of the cylinders 11. The engine 1 is amulti-cylinder engine.

A piston 3 is slidably inserted in each of the cylinders 11. The piston3 is coupled to a crankshaft 15 via a connecting rod 14. With thecylinder 11 and the cylinder head 13, the piston 3 defines thecombustion chamber 17. The term “combustion chamber” may be used in abroad sense. That is, the “combustion chamber” may mean the spacedefined by the piston 3, the cylinder 11, and the cylinder head 13regardless of a position of the piston 3.

A geometric compression ratio of the engine 1 is set to be equal to orhigher than 10 and equal to or lower than 30. As will be describedlater, in a part of an operating range, the engine 1 performs SPCCIcombustion in which spark ignition (SI) combustion and compressionignition (CI) combustion are combined. In the SPCCI combustion, the CIcombustion is controlled using heat and a pressure increase generated bythe SI combustion. The engine 1 is of a compression-ignition type. Inthis engine 1, a temperature of the combustion chamber 17 at the timewhen the piston 3 reaches compression top dead center (that is, acompression end temperature) does not have to be increased. Thegeometric compression ratio of the engine 1 can be set relatively low. Areduction in the geometric compression ratio is advantageous in terms ofa reduction in cooling loss and a reduction in mechanical loss. As anengine having a regular specification (using low octane fuel, an octanerating of which is approximately 91), the engine 1 may have thegeometric compression ratio of 14 to 17. As an engine having ahigh-octane specification (using high octane fuel, the octane rating ofwhich is approximately 96), the engine 1 may have the geometriccompression ratio of 15 to 18.

In the cylinder head 13, an intake port 18 is formed for each of thecylinders 11. Although not illustrated, the intake port 18 has a firstintake port and a second intake port. The intake port 18 communicateswith the combustion chamber 17. Although detailed illustration is notprovided, the intake port 18 is a so-called tumble port. That is, theintake port 18 has such a shape that a tumble flow is generated in thecombustion chamber 17.

An intake valve 21 is disposed in the intake port 18. The intake valve21 is opened/closed at a position between the combustion chamber 17 andthe intake port 18. The intake valve 21 is opened/closed at a specifiedtiming by a valve mechanism. The valve mechanism is preferably avariable valve mechanism that varies valve timing and/or valve lifting.In this configuration example, as illustrated in FIG. 2, the variablevalve mechanism includes an intake electric sequential-valve timing(S-VT) 23. The intake electric S-VT 23 continuously varies a rotationphase of an intake camshaft within a specified angle range. The intakevalve mechanism may include a hydraulic S-VT instead of the electricS-VT.

In the cylinder head 13, an exhaust port 19 is also formed for each ofthe cylinders 11. The exhaust port 19 also has a first exhaust port anda second exhaust port. The exhaust port 19 communicates with thecombustion chamber 17.

An exhaust valve 22 is disposed in the exhaust port 19. The exhaustvalve 22 is opened/closed at a position between the combustion chamber17 and the exhaust port 19. The exhaust valve 22 is opened/closed at aspecified timing by a valve mechanism. This valve mechanism ispreferably a variable valve mechanism that varies valve timing and/orvalve lifting. In this embodiment, as illustrated in FIG. 2, thevariable valve mechanism includes an exhaust electric S-VT 24. Theexhaust electric S-VT 24 continuously varies a rotation phase of anexhaust camshaft within a specified angle range. The exhaust valvemechanism may include a hydraulic S-VT instead of the electric S-VT.

The intake electric S-VT 23 and the exhaust electric S-VT 24 regulateduration of an overlap period in which both of the intake valve 21 andthe exhaust valve 22 are opened. In the case where the duration of theoverlap period is extended, residual gas in the combustion chamber 17can be eliminated. In addition, when the duration of the overlap periodis regulated, internal exhaust gas recirculation (EGR) gas can beintroduced into the combustion chamber 17. The intake electric S-VT 23and the exhaust electric S-VT 24 constitute an internal EGR system.Here, the internal EGR system is not limited to one constructed of theS-VTs.

For each of the cylinders 11, an injector 6 is attached to the cylinderhead 13. The injector 6 directly injects the fuel into the combustionchamber 17. The injector 6 is an example of a fuel supplier. Althoughdetailed illustration is not provided, the injector 6 is formed of afuel injection valve of a multiple injection-port type that has multipleinjection ports. The injector 6 is attached to a central portion of thecombustion chamber 17 and injects the fuel such that a fuel sprayspreads radially from a center of the combustion chamber 17.

A fuel supply system 61 is connected to the injector 6. The fuel supplysystem 61 includes: a fuel tank 63 configured to store the fuel; and afuel supply passage 62 that couples the fuel tank 63 and the injector 6to each other. A fuel pump 65 and a common rail 64 are provided in thefuel supply passage 62. The fuel pump 65 pumps the fuel to the commonrail 64. In the engine 1 of this configuration example, the fuel pump 65is a plunger pump that is driven by the crankshaft 15. The common rail64 stores the fuel, which is pumped from the fuel pump 65, at a highfuel pressure. When the injector 6 is opened, the fuel stored in thecommon rail 64 is injected into the combustion chamber 17 from theinjection ports of the injector 6. The fuel supply system 61 can supplythe high-pressure fuel at 30 MPa or higher to the injector 6. Thepressure of the fuel to be supplied to the injector 6 may vary accordingto an operation state of the engine 1. A configuration of the fuelsupply system 61 is not limited to the configuration described above.

For each of the cylinders 11, an ignition plug 25 is attached to thecylinder head 13. The ignition plug 25 forcibly ignites an air-fuelmixture in the combustion chamber 17. Although detailed illustration isnot provided, an electrode of the ignition plug 25 faces the inside ofthe combustion chamber 17 and is located near a ceiling surface of thecombustion chamber 17.

An intake passage 40 is connected to one side surface of the engine 1.The intake passage 40 communicates with the intake port 18 for each ofthe cylinders 11. Gas to be introduced into the combustion chamber 17flows through the intake passage 40. An air cleaner 41 is disposed in anupstream end portion of the intake passage 40. The air cleaner 41filters fresh air. A surge tank 42 is disposed near a downstream end ofthe intake passage 40. A portion of the intake passage 40 on adownstream side of the surge tank 42 constitutes an independent passagethat is branched for each of the cylinders 11. A downstream end of theindependent passage is connected to the intake port 18 for each of thecylinders 11.

A throttle valve 43 is disposed between the air cleaner 41 and the surgetank 42 in the intake passage 40. The throttle valve 43 regulates anopening degree thereof so as to regulate an introduction amount of thefresh air into the combustion chamber 17.

In addition, in the intake passage 40, a supercharger 44 is disposed ona downstream side of the throttle valve 43. The supercharger 44supercharges the gas to be introduced into the combustion chamber 17. Inthe engine 1 of this configuration example, the supercharger 44 is amechanical supercharger that is driven by the engine 1. The mechanicalsupercharger 44 may be of a root type, a Lysholm type, a vane type, or acentrifugal type.

An electromagnetic clutch 45 is interposed between the supercharger 44and the engine 1. At a position between the supercharger 44 and theengine 1, the electromagnetic clutch 45 transmits drive power from theengine 1 to the supercharger 44 and blocks the transmission of the drivepower. When an engine control unit (ECU) 10 switches between engagementand disengagement of the electromagnetic clutch 45, the supercharger 44is switched between ON and OFF.

In the intake passage 40, an intercooler 46 is disposed on a downstreamside of the supercharger 44. The intercooler 46 cools the gas that hasbeen compressed by the supercharger 44. The intercooler 46 may be of awater-cooling type or an oil-cooling type, for example.

A bypass passage 47 is connected to the intake passage 40. The bypasspassage 47 connects a portion of the intake passage 40 on an upstreamside of the supercharger 44 and a portion of the intake passage 40 on adownstream side of the intercooler 46, so as to bypass the supercharger44 and the intercooler 46. An air bypass valve 48 is disposed in thebypass passage 47. The air bypass valve 48 regulates a flow rate of thegas that flows through the bypass passage 47.

When the supercharger 44 is turned OFF (that is, when theelectromagnetic clutch 45 is disengaged), the ECU 10 fully opens the airbypass valve 48. The gas that flows through the intake passage 40bypasses the supercharger 44 and is introduced into the combustionchamber 17 of the engine 1. The engine 1 is operated in anon-supercharged state, that is, a naturally aspired state.

When the supercharger 44 is turned ON, the engine 1 is operated in asupercharged state. When the supercharger 44 is turned ON (that is, whenthe electromagnetic clutch 45 is engaged), the ECU 10 regulates anopening degree of the air bypass valve 48. Some of the gas that hasflowed through the supercharger 44 flows through the bypass passage 47and flows back to the upstream side of the supercharger 44. When the ECU10 regulates the opening degree of the air bypass valve 48, a boostpressure of the gas to be introduced into the combustion chamber 17varies. Here, a supercharged period may be defined as a period in whicha pressure in the surge tank 42 exceeds the atmospheric pressure, and anon-supercharged period may be defined as a period in which the pressurein the surge tank 42 becomes equal to or lower than the atmosphericpressure.

In the engine 1 of this configuration example, a supercharging system 49is constructed of the supercharger 44, the bypass passage 47, and theair bypass valve 48. For example, the supercharger 44 is turned OFF in alow-load, low-speed operating range and is turned ON in the operatingrange other than the above.

The engine 1 includes a swirl generation part, in which a swirl flow isgenerated, in the combustion chamber 17. The swirl generation partincludes a swirl control valve 56 that is attached to the intake passage40. Although detailed illustration is not provided, the swirl controlvalve 56 is disposed in a secondary passage among a primary passageconnected to the first intake port of the two intake ports 18 and thesecondary passage connected to the second intake port. The swirl controlvalve 56 is an opening-degree regulation valve that can reduce a crosssection of the secondary passage. When the opening degree of the swirlcontrol valve 56 is small, an intake flow rate from the first intakeport into the combustion chamber 17 is relatively large while the intakeflow rate from the second intake port into the combustion chamber 17 isrelatively small. As a result, the swirl flow in the combustion chamber17 is intensified. Meanwhile, when the opening degree of the swirlcontrol valve 56 is large, the intake flow rate from the first intakeport into the combustion chamber 17 and the intake flow rate from thesecond intake port into the combustion chamber 17 become substantiallythe same. As a result, the swirl flow in the combustion chamber 17 isweakened. When the swirl control valve 56 is fully opened, the swirlflow is not generated.

An exhaust passage 50 is connected to another side surface of the engine1. The exhaust passage 50 communicates with the exhaust port 19 for eachof the cylinders 11. The exhaust passage 50 is a passage through whichexhaust gas discharged from the combustion chamber 17 flows. Althoughdetailed illustration is not provided, an upstream portion of theexhaust passage 50 constitutes an independent passage that is branchedfor each of the cylinders 11. An upstream end of the independent passageis connected to the exhaust port 19 for each of the cylinders 11.

An exhaust gas aftertreatment system having a plurality of catalyticconverters is disposed in the exhaust passage 50. Although notillustrated, an upstream catalytic converter is disposed in an engineroom. The upstream catalytic converter has a three-way catalyst 511 anda gasoline particulate filter (GPF) 512. A downstream catalyticconverter is also disposed in the engine room. The downstream catalyticconverter has a three-way catalyst 513. The configuration of the exhaustgas aftertreatment system is not limited to the illustrated example ofthe configuration. For example, the GPF 512 may not be provided. Inaddition, the catalytic converter is not limited to one having thethree-way catalysts. Furthermore, an arrangement order of the three-waycatalysts and the GPF may appropriately be changed.

An EGR passage 52 that constitutes an external EGR system is connectedbetween the intake passage 40 and the exhaust passage 50. The EGRpassage 52 is a passage through which the exhaust gas is partiallyrecirculated into the intake passage 40. An upstream end of the EGRpassage 52 is connected to a portion of the exhaust passage 50 on adownstream side of the downstream catalytic converter. A downstream endof the EGR passage 52 is connected to the portion of the intake passage40 on the upstream side of the supercharger 44.

An EGR cooler 53 of a water-cooling type is disposed in the EGR passage52. The EGR cooler 53 cools the exhaust gas. An EGR valve 54 is alsodisposed in the EGR passage 52. The EGR valve 54 regulates a flow rateof the exhaust gas flowing through the EGR passage 52. When an openingdegree of the EGR valve 54 is regulated, a recirculation amount of thecooled exhaust gas, that is, external EGR gas can be regulated.

In the engine 1 of this configuration example, an EGR system 55 isconstructed of the external EGR system and the internal EGR system. Theexternal EGR system can supply the exhaust gas, the temperature of whichis lower than that in the internal EGR system, to the combustion chamber17.

The engine 1 has a cooling system 70. The cooling system 70 includes: apump 71 that supplies a coolant; an inlet passage 72 through which thecoolant flows from the pump 71 into the cylinder block 12 of the engine1; a radiator passage 74 through which the coolant, which has flowedthrough a coolant passage in the engine 1, flows from the cylinder head13 of the engine 1 into the pump 71 via a radiator 73; and aradiator-bypass passage 75 through which the coolant, which has flowedthrough the coolant passage in the engine 1, bypasses the radiator 73and flows into the pump 71.

The pump 71 is a mechanical pump that is driven in an interlockingmanner with the crankshaft 15. A discharge port of the pump 71 isconnected to the inlet passage 72. The pump 71 is provided with a firstcoolant temperature sensor SW9 that detects a temperature of the coolantto be discharged to the inlet passage 72. A discharge amount of thecoolant from the pump 71 fluctuates according to an engine speed and arecirculation amount of the coolant into the pump 71. The first coolanttemperature sensor SW9 may be arranged in a manner to detect thetemperature of the coolant flowing through the inlet passage 72.

The inlet passage 72 communicates between the discharge port of the pump71 and an inlet of the coolant passage in the cylinder block 12.

The radiator passage 74 is connected to the coolant passage in thecylinder head 13. In the radiator passage 74, a thermostat valve 76 isarranged between the radiator 73 and the pump 71. The thermostat valve76 is formed of an electric thermostat valve. More specifically, thethermostat valve 76 is a general thermostat valve having a heating wiretherein. The thermostat valve 76 is configured to be opened according tothe temperature of the coolant when the temperature of the coolant isequal to or higher than a specified coolant temperature during ade-energized period. However, with energization of the heating wire, thethermostat valve 76 can be opened even when the temperature of thecoolant is lower than the specified coolant temperature. That is, in thede-energized period, the thermostat valve 76 is opened at the specifiedcoolant temperature, and thus, the temperature of the coolant in theradiator passage 74 can be brought closer to the specified coolanttemperature. Meanwhile, in the energized period, the thermostat valve 76is opened at a desired coolant temperature that is lower than thespecified coolant temperature. Accordingly, the temperature of thecoolant in the radiator passage 74 can be brought to the desired coolanttemperature.

Similar to the radiator passage 74, the radiator-bypass passage 75 isconnected to the coolant passage in the cylinder head 13. A flow rateregulator valve 77 is arranged in an intermediate portion of theradiator-bypass passage 75. The flow rate regulator valve 77 is anon/off-type valve that can be switched between an open state at aspecified opening degree and a closed state of being fully closed. Theflow rate regulator valve 77 regulates a period in the open state and aperiod in the closed state, more specifically, a ratio between the openstate and the closed state per unit time (hereinafter referred to as aduty ratio), so as to regulate the flow rate of the coolant flowingthrough the radiator-bypass passage 75.

(Configuration of Engine Controller)

The engine controller includes the ECU 10. The ECU 10 is a controllerthat has a well-known microcomputer as a base, and, as illustrated inFIG. 2, includes a processor (e.g., a central processing unit (CPU)) 101that executed programs, memory 102 constructed of random access memory(RAM) or read only memory (ROM), for example, to store a program anddata, and an input/output bus 103 that inputs/outputs an electricsignal. The ECU 10 is an example of a control unit.

As illustrated in FIG. 1 and FIG. 2, various sensors SW1 to SW17 areconnected to the ECU 10. Each of the sensors SW1 to SW17 outputs asignal to the ECU 10. The following sensors are included.

Airflow sensor SW1: arranged on a downstream side of the air cleaner 41in the intake passage 40 to measure the flow rate of the fresh airflowing through the intake passage 40.

First intake temperature sensor SW2: arranged on the downstream side ofthe air cleaner 41 in the intake passage 40 so as to measure atemperature of the fresh air flowing through the intake passage 40.

First pressure sensor SW3: arranged on a downstream side of a position,to which the EGR passage 52 is connected, in the intake passage 40 andon the upstream side of the supercharger 44 so as to measure a pressureof the gas flowing into the supercharger 44.

Second intake temperature sensor SW4: arranged on the downstream side ofthe supercharger 44 and on the upstream side of a position, to which thebypass passage 47 is connected, in the intake passage 40 so as tomeasure the temperature of the gas that has flowed out of thesupercharger 44.

Second pressure sensor SW5: attached to the surge tank 42 to measure thepressure of the gas on the downstream side of the supercharger 44.

In-cylinder pressure sensor SW6: attached to the cylinder head 13 in amanner to correspond to each of the cylinders 11 so as to measure apressure in each of the combustion chambers 17.

Linear O₂ sensor SW7: arranged on an upstream side of the upstreamcatalytic converter in the exhaust passage 50 so as to measureconcentration of oxygen in the exhaust gas.

Lambda O₂ sensor SW8: arranged on a downstream side of the three-waycatalyst 511 in the upstream catalytic converter so as to measure theconcentration of oxygen in the exhaust gas.

First coolant temperature sensor SW9: as described above, attached tothe pump 71 to detect the temperature of the coolant flowing into thecylinder block 12.

Second coolant temperature sensor SW10: attached to the engine 1 tomeasure the temperature of the coolant immediately after heat exchangewith the engine 1.

Crank angle sensor SW11: attached to the engine 1 to measure a rotationangle of the crankshaft 15.

Accelerator operation amount sensor SW12: attached to an acceleratorpedal mechanism to measure an accelerator operation amount thatcorresponds to an operation amount of an accelerator pedal.

Intake cam angle sensor SW13: attached to the engine 1 to measure arotation angle of the intake camshaft.

Exhaust cam angle sensor SW14: attached to the engine 1 to measure arotation angle of the exhaust camshaft.

EGR differential pressure sensor SW15: arranged in the EGR passage 52 tomeasure a differential pressure between the upstream side and thedownstream side of the EGR valve 54.

Fuel pressure sensor SW16: attached to the common rail 64 in the fuelsupply system 61 to measure the pressure of the fuel to be supplied tothe injector 6.

Third intake temperature sensor SW17: attached to the surge tank 42 tomeasure the temperature of the gas in the surge tank 42, in other words,the temperature of the intake air to be introduced into the combustionchamber 17.

On the basis of signals from these sensors SW1 to SW17, the ECU 10determines the operation state of the engine 1 and calculates a controlamount of each of the devices according to a predetermined controllogic. The control logic is stored in the memory 102. The control logicincludes calculation of a target amount and/or the control amount byusing an operation map stored in the memory 102.

The ECU 10 outputs an electric signal related to the calculated controlamount to the injector 6, the ignition plug 25, the intake electric S-VT23, the exhaust electric S-VT 24, the fuel supply system 61, thethrottle valve 43, the EGR valve 54, the electromagnetic clutch 45 ofthe supercharger 44, the air bypass valve 48, the swirl control valve56, the thermostat valve 76, and the flow rate regulator valve 77.

For example, based on the signal of the accelerator operation amountsensor SW12 and the operation map, the ECU 10 sets target torque of theengine 1 and determines a target boost pressure. Then, based on thetarget boost pressure and the differential pressure before and after thesupercharger 44 acquired from the signals of the first pressure sensorSW3 and the second pressure sensor SW5, the ECU 10 executes a feedbackcontrol for regulating the opening degree of the air bypass valve 48, soas to bring the boost pressure to the target boost pressure.

In addition, the ECU 10 sets a target EGR rate on the basis of theoperation state of the engine 1 and the operation map. The EGR rate is arate of the EGR gas to the whole gas in the combustion chamber 17. TheECU 10 determines a target EGR gas amount on the basis of an intake airamount that is based on the target EGR rate and the signal of theaccelerator operation amount sensor SW12, and executes the feedbackcontrol for regulating the opening degree of the EGR valve 54 on thebasis of the differential pressure before and after the EGR valve 54acquired from the signal of the EGR differential pressure sensor SW15,so as to set an external EGR gas amount to be introduced into thecombustion chamber 17 to the target EGR gas amount.

Furthermore, the ECU 10 executes an air-fuel ratio feedback control inthe case where a specified control condition is established. Morespecifically, the ECU 10 regulates a fuel injection amount of theinjector 6 such that the air-fuel ratio of the air-fuel mixture acquiresa desired value on the basis of the concentration of oxygen in theexhaust gas measured by the linear O₂ sensor SW7 and the lambda O₂sensor SW8.

(Concept of SPCCI Combustion)

With improvement in fuel economy and improvement in exhaust gasperformance as primary purposes, the engine 1 performs combustion bycompression self-ignition in a specified operation state. In regard tothe combustion by the self-ignition, self-ignition timing variessignificantly when the temperature inside the combustion chamber 17before initiation of the compression fluctuates. Thus, the engine 1performs the SPCCI combustion in which the SI combustion and the CIcombustion are combined.

The SPCCI combustion is a mode in which the ignition plug 25 forciblyignites the air-fuel mixture in the combustion chamber 17 and the SIcombustion of the air-fuel mixture is performed by flame propagation andin which the CI combustion of unburned air-fuel mixture is performed bythe self-ignition when the heat generated by the SI combustion increasesthe temperature inside the combustion chamber 17 and the pressure in thecombustion chamber 17 is increased by the flame propagation.

When a heat generation amount by the SI combustion is regulated, thefluctuation of the temperature inside the combustion chamber 17 beforethe initiation of the compression can be offset. When the ECU 10regulates the ignition timing, the air-fuel mixture can be subjected tothe self-ignition at target timing.

In the SPCCI combustion, the heat generates at a lower rate in the SIcombustion than in the CI combustion. FIG. 3 exemplifies a waveform 301of a heat generation rate in the SPCCI combustion. The waveform 301 ofthe heat generation rate in the SPCCI combustion has a shallowergradient at initial rise than a gradient at initial rise of a waveformin the CI combustion. In addition, a pressure fluctuation (dp/dθ) in thecombustion chamber 17 is lower in the SI combustion than in the CIcombustion.

When the unburned air-fuel mixture is self-ignited after the initiationof the SI combustion, at the self-ignition timing, the gradient of thewaveform of the heat generation rate is possibly changed from beingshallow to being steep. The waveform of the heat generation ratepossibly has a flexion point X at timing θci at which the CI combustionis initiated.

After the initiation of the CI combustion, the SI combustion and the CIcombustion are performed in parallel. The CI combustion produces moreheat than the SI combustion. Thus, the heat generation rate in the CIcombustion is relatively higher than that in the SI combustion. However,the CI combustion is performed after the piston 3 reaches thecompression top dead center. Thus, the gradient of the waveform of theheat generation rate is avoided from becoming excessively steep. Thepressure fluctuation (dp/dθ) in the CI combustion is also relativelylow.

The pressure fluctuation (dp/dθ) can be used as an index that representscombustion noise. As described above, since the pressure fluctuation(dp/dθ) is low in the SPCCI combustion, combustion noise is avoided frombecoming excessively large. Combustion noise of the engine 1 issuppressed to be equal to or lower than a permissible level.

When the CI combustion is terminated, the SPCCI combustion isterminated. Compared to a case of the SI combustion only that isexemplified by a waveform 302 in FIG. 3, a combustion period of the CIcombustion is short. Combustion termination timing in the SPCCIcombustion is earlier than that in the SI combustion.

(Engine Operating Range)

FIG. 4 exemplifies an operation map 401 related to control of the engine1. The operation map 401 is stored in the memory 102 of the ECU 10. Theoperation map 401 is an operation map of the engine 1 in a warm period.

The operation map 401 is defined by a load and the speed of the engine1. As boundaries are exemplified by solid lines in FIG. 4, the operationmap 401 is divided into three ranges according to magnitudes of thespeed and the load. More specifically, the operation map 401 is dividedinto a low-load range A1 in a low-speed, middle-speed range, middle-loadand large-load ranges A2, A3, A4 in the low-speed, middle-speed range,and a high-speed range A5 at a speed N1 and higher.

Here, the low-speed range, the middle-speed range, and the high-speedrange may respectively be set as the low-speed range, the middle-speedrange, and the high-speed range in the case where the entire operatingrange of the engine 1 is divided into three ranges of the low-speedrange, the middle-speed range, and the high-speed range in asubstantially equal manner in a speed direction. Alternatively, thelow-load range, the middle-load range, and the high-load range mayrespectively be set as the low-load range, the middle-load range, andthe high-load range in the case where the entire operating range of theengine 1 is divided into three ranges of the low-load range, themiddle-load range, and the high-load range in a substantially equalmanner in a load direction.

The operation map 401 indicates a state of the air-fuel mixture and acombustion mode in each of the ranges. The engine 1 performs the SPCCIcombustion in the ranges A1, A2, A3, A4. The engine 1 performs the SIcombustion in the range A5. A detailed description will hereinafter bemade on the operation of the engine 1 in each of the ranges of theoperation map 401 in FIG. 4 with reference to fuel injection timing andthe ignition timing illustrated in FIG. 5. A horizontal axis in FIG. 5represents a crank angle. The reference numerals 601, 602, 603, 604,605, 606 in FIG. 5 respectively correspond to the operation states ofthe engine 1 indicated by the reference numerals 601, 602, 603, 604,605, 606 in the operation map 401 illustrated in FIG. 4.

(Engine Operation in Range A1)

When the engine 1 is operated in the range A1, the engine 1 performs theSPCCI combustion. The reference numeral 601 in FIG. 5 represents thefuel injection timing (the reference numerals 6011, 6012), the ignitiontiming (the reference numeral 6013), and a combustion waveform (that is,a waveform representing a change in the heat generation rate withrespect to the crank angle, the reference numeral 6014) at the time whenthe engine 1 is operated in the operation state 601 in the range A1 .The reference numeral 602 represents the fuel injection timing (thereference numerals 6021, 6022), the ignition timing (the referencenumeral 6023), and the combustion waveform (the reference numeral 6024)at the time when the engine 1 is operated in the operation state 602 inthe range A1 . The reference numeral 603 represents the fuel injectiontiming (the reference numerals 6031, 6032), the ignition timing (thereference numeral 6033), and the combustion waveform (the referencenumeral 6034) at the time when the engine 1 is operated in the operationstate 603 in the range A1. In the operation states 601, 602, 603, thespeed of the engine 1 is the same while the load thereof differs. In theoperation state 601, the load is the smallest (that is, the light load),which is followed by the operation state 602 (that is, the small load).Of these, the load is the largest in the operation state 603.

In order to improve fuel efficiency of the engine 1, the EGR system 55introduces the EGR gas into the combustion chamber 17. Morespecifically, the intake electric S-VT 23 and the exhaust electric S-VT24 are provided with a positive overlap period in which both of theintake valve 21 and the exhaust valve 22 are opened near exhaust topdead center. Some of the exhaust gas that is discharged from thecombustion chamber 17 to the intake port 18 and the exhaust port 19 isintroduced into the combustion chamber 17 again. Since the hot exhaustgas is introduced into the combustion chamber 17, the temperature insidethe combustion chamber 17 is increased. This is advantageous forstabilization of the SPCCI combustion. The intake electric S-VT 23 andthe exhaust electric S-VT 24 may be provided with a negative overlapperiod in which both of the intake valve 21 and the exhaust valve 22 areclosed.

The swirl generation part generates the strong swirl flow in thecombustion chamber 17. A swirl ratio is equal to or higher than 4, forexample. The swirl control valve 56 is fully closed or at a specifiedopening degree on a closing side. As described above, since the intakeport 18 is the tumble port, an oblique swirl flow having a tumblecomponent and a swirl component is generated in the combustion chamber17.

In the intake stroke, the injector 6 injects the fuel into thecombustion chamber 17 for multiple times (the reference numerals 6011,6012, 6021, 6022, 6031, 6032). The air-fuel mixture is stratified by themultiple times of the fuel injection and the swirl flow in thecombustion chamber 17.

Concentration of the fuel in the air-fuel mixture in the central portionof the combustion chamber 17 is higher than the concentration of thefuel therein in an outer circumferential portion. More specifically, theair-fuel ratio (A/F) of the air-fuel mixture in the central portion isequal to or higher than 20 and equal to or lower than 30, and the A/F ofthe air-fuel mixture in the outer circumferential portion is equal to orhigher than 35. Here, a value of the air-fuel ratio is a value of theair-fuel ratio at the time of the ignition, and the same applies to thefollowing description. When the A/F of the air-fuel mixture near theignition plug 25 is set to be equal to or higher than 20 and equal to orlower than 30, it is possible to suppress generation of raw NO_(x)during the SI combustion. In addition, when the A/F of the air-fuelmixture in the outer circumferential portion is set to be equal to orhigher than 35, the CI combustion is stabilized.

The A/F of the air-fuel mixture is leaner than the stoichiometricair-fuel ratio in the entire combustion chamber 17 (that is, an excessair ratio λ>1). In detail, the A/F of the air-fuel mixture is equal toor higher than 30 in the entire combustion chamber 17. In this way, itis possible to suppress the generation of raw NO_(x) and thus to improvethe exhaust gas performance.

When the load of the engine 1 is small (that is, at the time of theoperation state 601), the injector 6 performs the first injection 6011in a first half of the intake stroke and performs the second injection6012 in a second half of the intake stroke. The first half of the intakestroke may be a first half when the intake stroke is equally dividedinto the first half and the second half, and the second half of theintake stroke may be the second half when the intake stroke is equallydivided into two. An injection amount ratio between the first injection6011 and the second injection 6012 may be 9:1, for example.

When the engine 1 is in the operation state 602 with the large load, theinjector 6 initiates the second injection 6022, which is performed inthe second half of the intake stroke, at an advanced timing from thetiming of the second injection 6012 in the operation state 601. Sincethe second injection 6022 is advanced, the air-fuel mixture in thecombustion chamber 17 becomes nearly homogenized. The injection amountratio between first injection 6021 and the second injection 6022 may beset to 7:3 to 8:2, for example.

When the engine 1 is in the operation state 603 with the largest load,the injector 6 initiates the second injection 6032, which is performedin the second half of the intake stroke, at a more advanced timing fromthe timing of the second injection 6022 in the operation state 602.Since the second injection 6032 is advanced further, the air-fuelmixture in the combustion chamber 17 becomes further nearly homogenized.The injection amount ratio between the first injection 6031 and thesecond injection 6032 may be set to 6:4, for example.

After the termination of the fuel injection, the ignition plug 25ignites the air-fuel mixture in the central portion of the combustionchamber 17 at the specified timing before compression top dead center(the reference numerals 6013, 6023, 6033). The ignition timing may beset at a termination period of the compression stroke. The terminationperiod of the compression stroke may be set as the termination periodwhen the compression stroke is equally divided into three periods of aninitiation period, a middle period, and the termination period.

As described above, since the air-fuel mixture in the central portionhas the relatively high concentration of the fuel, ignitability isimproved, and the SI combustion by the flame propagation is stabilized.When the SI combustion is stabilized, the CI combustion is initiated atan appropriate timing. As a result, controllability of the CI combustionis improved. In addition, the generation of combustion noise issuppressed. Furthermore, when the SPCCI combustion is performed bysetting the A/F of the air-fuel mixture to be lean, fuel efficiency ofthe engine 1 can be improved significantly.

(Engine Operation in Ranges A2, A3, A4)

In the case where the engine 1 is operated in any of the ranges A2, A3,A4, the engine 1 performs the SPCCI combustion. The reference numeral604 in FIG. 5 represents the fuel injection timing (the referencenumerals 6041, 6042), the ignition timing (the reference numeral 6043),and the combustion waveform (the reference numeral 6044) at the timewhen the engine 1 is operated in the operation state 604 in the rangeA2. The reference numeral 605 represents the fuel injection timing (thereference numeral 6051), the ignition timing (the reference numeral6052), and the combustion waveform (the reference numeral 6053) at thetime when the engine 1 is operated in the operation state 605 in therange A4.

The EGR system 55 introduces the EGR gas into the combustion chamber 17.More specifically, the intake electric S-VT 23 and the exhaust electricS-VT 24 are provided with the positive overlap period in which both ofthe intake valve 21 and the exhaust valve 22 are opened near the exhausttop dead center. Internal EGR gas is introduced into the combustionchamber 17. In addition, the EGR system 55 introduces the exhaust gas,which is cooled by the EGR cooler 53, into the combustion chamber 17through the EGR passage 52. That is, the external EGR gas, a temperatureof which is lower than that of the internal EGR gas, is introduced intothe combustion chamber 17. The external EGR gas regulates thetemperature inside the combustion chamber 17 to an appropriatetemperature. The EGR system 55 reduces the EGR gas amount according tothe increase in the load of the engine 1. With a full-open load, the EGRsystem 55 may reduce the EGR gas including the internal EGR gas and theexternal EGR gas to zero.

In the range A2 and the range A3, the swirl control valve 56 is fullyclosed or at a specified opening degree on a closing side. The strongswirl flow, the swirl ratio of which is equal to or higher than 4, isgenerated in the combustion chamber 17. Meanwhile, in the range A4, theswirl control valve 56 is opened.

The A/F of the air-fuel mixture is the stoichiometric air-fuel ratio inthe entire combustion chamber 17 (A/F≈14.7). The three-way catalysts511, 513 purify the exhaust gas that has been discharged from thecombustion chamber 17. Thus, the exhaust gas performance of the engine 1becomes favorable. The A/F of the air-fuel mixture only needs to fallwithin a purification window of the three-way catalysts. The excess airratio λ of the air-fuel mixture may be set to 1.0±0.2. When the engine 1is operated in the range A3 including the maximum load, the A/F of theair-fuel mixture may be set to the stoichiometric air-fuel ratio or bericher than the stoichiometric air-fuel ratio in the entire combustionchamber 17 (that is, the excess air ratio λ of the air-fuel mixture isλ≤1).

Since the EGR gas is introduced into the combustion chamber 17, agas-fuel ratio (G/F) as a weight ratio between the whole gas and thefuel in the combustion chamber 17 is leaner than the stoichiometricair-fuel ratio. The G/F of the air-fuel mixture may be equal to orhigher than 18. In this way, occurrence of so-called knocking can beavoided. The G/F may be set to be equal to or higher than 18 and equalto or lower than 30. The G/F may be set to be equal to or higher than 18and equal to or lower than 50.

When the engine 1 is operated in the operation state 604, the injector 6injects the fuel for the multiple times (the reference numerals 6041,6042) during the intake stroke. The injector 6 may perform the firstinjection 6041 in the first half of the intake stroke and the secondinjection 6042 in the second half of the intake stroke.

When the engine 1 is operated in the operation state 605, the injector 6injects the fuel (the reference numeral 6051) during the intake stroke.

After the fuel is injected, the ignition plug 25 ignites the air-fuelmixture at the specified timing near the compression top dead center(the reference numerals 6043, 6052). When the engine 1 is operated inthe operation state 604, the ignition plug 25 may ignite the air-fuelmixture before the compression top dead center (the reference numeral6043). When the engine 1 is operated in the operation state 605, theignition plug 25 may ignite the air-fuel mixture after the compressiontop dead center (the reference numeral 6052).

The SPCCI combustion is performed by setting the A/F of the air-fuelmixture to the stoichiometric air-fuel ratio. In this way, the exhaustgas discharged from the combustion chamber 17 can be purified by usingthe three-way catalysts 511, 513. In addition, when the EGR gas isintroduced into the combustion chamber 17 to dilute the air-fuelmixture, fuel efficiency of the engine 1 is improved.

(Engine Operation in Range A5)

When the speed of the engine 1 is high, a time required to change thecrank angle by 1° is shortened. When the speed of the engine 1 is high,it is difficult to stratify the air-fuel mixture in the combustionchamber 17, which makes it difficult to perform the SPCCI combustion.Thus, in the case where the engine 1 is operated in the range A5, theengine 1 performs the SI combustion instead of the SPCCI combustion.

The reference numeral 606 in FIG. 5 represents the fuel injection timing(the reference numeral 6061), the ignition timing (the reference numeral6062), and the combustion waveform (the reference numeral 6063) at thetime when the engine 1 is operated in the operation state 606 with thelarge load in the range A5.

The EGR system 55 introduces the EGR gas into the combustion chamber 17.The EGR system 55 reduces the amount of the EGR gas according to theincrease in the load. With the full-open load, the EGR system 55 mayreduce the EGR gas to zero.

The swirl control valve 56 is fully opened. In the combustion chamber17, the swirl flow is not generated, and only the tumble flow isgenerated. When the swirl control valve 56 is fully opened, chargingefficiency can be improved, and pumping loss can be reduced.

The A/F of the air-fuel mixture is basically the stoichiometric air-fuelratio in the entire combustion chamber 17 (A/F≈14.7). The excess airratio λ of the air-fuel mixture is preferably set to 1.0±0.2. When theengine 1 is operated with the load near the full-open load, the excessair ratio λ of the air-fuel mixture may be lower than 1.

The injector 6 initiates the fuel injection during the intake stroke.The injector 6 injects the fuel all at once (the reference numeral6061). Since the fuel injection is initiated during the intake stroke, ahomogenous or a substantially homogenous air-fuel mixture is produced inthe combustion chamber 17. In addition, since a long fuel vaporizationtime can be secured, it is possible to reduce unburned fuel loss.

After the fuel injection is terminated, the ignition plug 25 ignites theair-fuel mixture at an appropriate timing before the compression topdead center (the reference numeral 6062).

(Estimation of Temperature of Exhaust Gas)

The temperature of the exhaust gas is used for various applications inthe control of the engine 1. For example, in the case where atemperature of the catalyst in the exhaust gas aftertreatment system isexcessively high, reliability of the catalyst is degraded. Thetemperature of the exhaust gas flowing into the catalyst is monitored.Then, in the case where the temperature of the exhaust gas isexcessively high, the temperature of the exhaust gas has to be reduced.In addition, for example, in order to regulate a recirculation amount ofthe EGR gas, the ECU 10 has to comprehend the temperature of the exhaustgas.

Attachment of a temperature sensor to the exhaust passage 50 of theengine 1 increases cost. When a plurality of sensors are used to measurethe temperature of the exhaust gas flowing into the catalyst and thetemperature of the exhaust gas flowing into the EGR passage 52, the costis further increased. Accordingly, the controller in this engine 1 isconfigured to estimate the temperature of the exhaust gas in anuppermost stream portion of the exhaust passage 50 on the basis of acombustion state of the air-fuel mixture in the cylinder withoutattaching the temperature sensor to the exhaust passage 50. Whenestimating the temperature of the exhaust gas in the uppermost streamportion of the exhaust passage 50, the controller can further estimatethe temperature of the exhaust gas flowing into the catalyst, thetemperature of the exhaust gas flowing into the EGR passage 52, and thelike on the basis of the temperature of the exhaust gas in the uppermoststream portion.

FIG. 6 exemplifies software modules of the ECU 10 stored in the memory102 which are executed by the processor 101 to perform their respectivefunctions related to the estimation of the temperature of the exhaustgas. The ECU 10 has an acquisition module 104, a calculation module 105,an estimation module 106, and a correction module 107.

The acquisition module 104 acquires the signals of the in-cylinderpressure sensor SW6 and the crank angle sensor SW11.

The calculation module 105 calculates progress of the combustion, whichis the crank angle at the time when the combustion in the cylinder isprogressed to a particular extent, on the basis of the signal of thein-cylinder pressure sensor SW6 and the signal of the crank angle sensorSW11 acquired by the acquisition module 104. More specifically, thecalculation module 105 may calculate the crank angle at which a masscombustion rate is 50%, that is, mfb50 (or a combustion gravity center).A method for calculating mfb50 may be a known method. More specifically,as exemplified in FIG. 3, the calculation module 105 can calculate mfb50from an area of the waveform 301 or 302 of the heat generation rate,which is acquired from the signal of the in-cylinder pressure sensor SW6and the signal of the crank angle sensor SW11. In the case where themass combustion rate is correlated with the temperature of the exhaustgas, a value of the mass combustion rate, which can be used as theprogress of the combustion, can be set to a value such as mfb10 ormfb90.

The calculation module 105 may calculate the combustion gravity centeras the progress of the combustion on the basis of the signal of thecrank angle sensor SW11, for example. For example, a waveform 303 inFIG. 3 exemplifies a temporal change in the signal of the crank anglesensor SW11. When the CI combustion is initiated during the SPCCIcombustion, a crank speed is increased. Thus, a pulse interval of thecrank angle sensor SW11 is reduced near the flexion point X. Inaddition, on the waveform 301 of the heat generation rate of the SPCCIcombustion, the flexion point X and mfb50 occur at the crank angles thatare close to each other. Thus, mfb50 may be calculated based on thepulse interval of the crank angle sensor SW11.

The estimation module 106 estimates the temperature of the exhaust gason the basis of a relationship between the progress of the combustionand the temperature of the exhaust gas. In detail, the estimation module106 uses a model representing the relationship between the progress ofthe combustion and the temperature of the exhaust gas, so as to estimatethe temperature of the exhaust gas from the progress of the combustioncalculated by the calculation module 105. The model is stored in thememory 102. More specifically, the model is configured as illustrated inFIG. 8, FIG. 9, and FIG. 11. A detailed description on the configurationof the model will be made later.

The correction module 107 corrects the temperature of the exhaust gas,which is estimated by the estimation module 106, according to thetemperature of the engine 1. A detailed description on the correction bythe correction module 107 will be made later.

Next, a description will be made on a procedure of estimating thetemperature of the exhaust gas executed by the ECU 10 with reference toa flowchart in FIG. 7. An order of steps in the flowchart illustrated inFIG. 7 can be changed.

In step S1, the ECU 10 reads the signal of each of the sensors SW1 toSW17. In following step S2, the ECU 10 injects and ignites the fuelaccording to the operation state of the engine 1 (see FIG. 4 and FIG.5).

Next, in step S3, the acquisition module 104 of the ECU 10 acquires thesignals of the in-cylinder pressure sensor SW6 and the crank anglesensor SW11 during the combustion in the cylinder. In following step S4,the calculation module 105 calculates the progress of the combustion.

In step S5, the estimation module 106 of the ECU 10 determines whether acombustion mode of the engine 1 is the SI combustion. If the combustionmode is the SI combustion, and thus, the determination in step S5 isYES, the processing proceeds to step S6. If the combustion mode is theSPCCI combustion and thus the determination in step S5 is NO, theprocessing proceeds to step S8.

In step S8, the estimation module 106 determines whether the air-fuelratio of the air-fuel mixture is the stoichiometric air-fuel ratio. Ifthe air-fuel ratio is the stoichiometric air-fuel ratio (that is, λ=1),and thus, the determination in step S8 is YES, the processing proceedsto step S9. If the air-fuel ratio is lean (that is, λ>1), and thus, thedetermination in step S8 is NO, the processing proceeds to step S11.

In step S6, the estimation module 106 estimates the temperature of theexhaust gas from the model stored in the memory 102 and the calculateddegree of the progress of the combustion. FIG. 8 exemplifies a modelthat represents the relationship between the progress of the combustionand the temperature of the exhaust gas. FIG. 8 exemplifies a model 801in the SI combustion, a model 802 in the SPCCI combustion at the timewhen the air-fuel ratio is the stoichiometric air-fuel ratio, and amodel 803 in the SPCCI combustion at the time when the air-fuel ratio islean.

As it is understood from FIG. 8, each of the models is represented by alinear function. This is because the present inventors have found that alinear correlation existed between the progress of the combustion andthe temperature of the exhaust gas in the case where the progress of thecombustion was delayed from the progress of the combustion at MBT.

In the SI combustion and the SPCCI combustion, the air-fuel mixture isignited. When the ignition timing is delayed, the temperature of theexhaust gas is increased. Thus, a correlation exists between theignition timing and the temperature of the exhaust gas. However,according to the research by the present inventors, in the so-calledspark-ignition engine in which the SI combustion and/or the SPCCIcombustion is performed, the relationship between the ignition timingand the temperature of the exhaust gas was not linear but non-linear.

When the ignition timing is delayed, the combustion is initiated in thepower stroke. The power stroke is a stroke in which a volume inside thecylinder is increased. The increase in the volume has a non-linearrelationship with advancement in the crank angle. The combustion in thepower stroke is affected by the non-linear increase in the volume. Thedegree of the progress of the combustion has a non-linear relationship(for example, a quadratic function) with a delay in the ignition timing.

The degree of the progress of the combustion is a parameter thatrepresents the combustion state. Thus, a linear correlation isestablished between the progress of the combustion and the illustratedwork of the engine 1. In addition, the temperature of the exhaust gas isdetermined from the quantity of heat, which is acquired by subtractingthe quantity of heat used for the illustrated work of the engine fromthe quantity of heat generated by the combustion in the cylinder.Accordingly, a linear correlation is also established between theillustrated work of the engine and the temperature of the exhaust gas.Since the ignition timing and the progress of the combustion have thenon-linear relationship, the correlation between the ignition timing andthe temperature of the exhaust gas is non-linear. Meanwhile, thecorrelation between the progress of the combustion and the temperatureof the exhaust gas is linear.

Thus, each of the models 801, 802, 803 is configured such that, in thecase where the progress of the combustion is delayed from the progressof the combustion at maximum brake torque (MBT), the temperature of theexhaust gas is estimated to be higher in a linear manner as the progressof the combustion is delayed. In the case where the progress of thecombustion is delayed from the progress of the combustion at the MBT,the estimation module 106 estimates the temperature of the exhaust gasto be higher in the linear manner as the progress of the combustion isdelayed according to the models 801, 802, 803.

Meanwhile, in the case where the ignition timing is excessivelyadvanced, abnormal combustion possibly occurs. As an advancement limitof the ignition timing at which the abnormal combustion does not occur,a knock limit is set. The knock limit is set on a delayed side from theMBT. The ignition timing is set to be delayed from the knock limit.

As illustrated in FIG. 8, even in the case where the progress of thecombustion is delayed from the progress of the combustion at the knocklimit, the linear correlation is established between the progress of thecombustion and the temperature of the exhaust gas. Each of the models801, 802, 803 is configured such that, in the case where the progress ofthe combustion is delayed from the progress of the combustion at theknock limit, the temperature of the exhaust gas is estimated to behigher in the linear manner as the progress of the combustion isdelayed. In the case where the progress of the combustion is delayedfrom the progress of the combustion at the knock limit, the estimationmodule 106 estimates the temperature of the exhaust gas to be higher inthe linear manner as the progress of the combustion is delayed accordingto the models 801, 802, 803.

Each of the models 801, 802, 803 is linear. Thus, even in the case wherethe progress of the combustion is significantly delayed from theprogress of the combustion at the MBT, the estimation module 106 canaccurately estimate the temperature of the exhaust gas.

The model 801 in the SI combustion differs from the models 802, 803 inthe SPCCI combustion. The model 801 represents a first relationshipbetween the progress of the combustion and the temperature of theexhaust gas in the SI combustion. Each of the models 802, 803 representsa second relationship between the progress of the combustion and thetemperature of the exhaust gas in the SPCCI combustion. In response toswitching of the combustion mode, the estimation module 106 switches themodel among the models 801, 802, 803 and estimates the temperature ofthe exhaust gas. In this way, in each of the SI combustion and the SPCCIcombustion, the estimation module 106 can accurately estimate thetemperature of the exhaust gas.

In detail, the model 801 in the SI combustion is configured such thatthe temperature of the exhaust gas at the same degree of the progress ofthe combustion is estimated to be higher than that in each of the models802, 803 in the SPCCI combustion. Thus, in the case where the progressof the combustion is the same, the estimation module 106 estimates thetemperature of the exhaust gas to be higher in the SI combustion than inthe SPCCI combustion.

In the SPCCI combustion, some of the unburned air-fuel mixture is burnedby the self-ignition. Thus, thermal efficiency in the SPCCI combustionis higher than that in the SI combustion. As a result, the temperatureof the exhaust gas is lower in the SPCCI combustion than in the SIcombustion. On the contrary, the temperature of the exhaust gas ishigher in the SI combustion than in the SPCCI combustion. The model 801in the SI combustion is configured such that the temperature of theexhaust gas at the same degree of the progress of the combustion isestimated to be higher than that in each of the models 802, 803 in theSPCCI combustion. Thus, in each of the SI combustion and the SPCCIcombustion, the estimation module 106 can accurately estimate thetemperature of the exhaust gas.

The model 802 in the SPCCI combustion is configured such that, when theair-fuel ratio of the air-fuel mixture is the stoichiometric air-fuelratio, a temperature increasing rate of the exhaust gas with respect toa change in the progress of the combustion is higher than thetemperature increasing rate of the model 801 in the SI combustion. Agradient of each of the straight lines in FIG. 8 corresponds to a“temperature increasing rate of the exhaust gas with respect to thechange in the progress of the combustion”.

As described above, thermal efficiency is high in the SPCCI combustion.Meanwhile, in the SPCCI combustion, the air-fuel mixture is partiallysubjected to the CI combustion. The CI combustion in the power stroke issignificantly affected by a change in the volume. More specifically,when the progress of the combustion is delayed, thermal efficiency inthe SPCCI combustion is significantly degraded. In the SPCCI combustion,the temperature of the exhaust gas is significantly changed with respectto the change in the progress of the combustion. Meanwhile, the SIcombustion by the flame propagation is less likely to be affected by thechange in the volume. Even when the progress of the combustion isdelayed in the SI combustion, thermal efficiency in the SI combustion isnot significantly degraded. In the SI combustion, the change in thetemperature of the exhaust gas with respect to the change in theprogress of the combustion is small.

Thus, in the model 802 in the SPCCI combustion, the temperatureincreasing rate of the exhaust gas with respect to the change in theprogress of the combustion is set to be higher than that in the model801 in the SI combustion. That is, the gradient of the model 802 issteeper than that of the model 801. In this way, in each of the SIcombustion and the SPCCI combustion, the estimation module 106 canaccurately estimate the temperature of the exhaust gas.

Each of the models 801, 802 at the time when the air-fuel ratio of theair-fuel mixture is stoichiometric air-fuel ratio differs from the model803 at the time when the air-fuel ratio of the air-fuel mixture is lean.Furthermore, in the SPCCI combustion, each of the models 801, 802 at thetime when the air-fuel ratio of the air-fuel mixture is thestoichiometric air-fuel ratio differs from the model 803 at the timewhen the air-fuel ratio of the air-fuel mixture is lean. The estimationmodule 106 switches between the models 802, 803 in response to switchingof the air-fuel ratio of the air-fuel mixture in the SPCCI combustion,so as to estimate the temperature of the exhaust gas. Thus, theestimation module 106 can accurately estimate the temperature of theexhaust gas.

In detail, the model 803 at the time when the air-fuel ratio of theair-fuel mixture is lean is configured such that the temperature of theexhaust gas at the same degree of the progress of the combustion isestimated to be lower than that in the model 802 at the time when theair-fuel ratio is the stoichiometric air-fuel ratio. Thus, in the casewhere the progress of the combustion is the same, the estimation module106 estimates the temperature of the exhaust gas to be lower at the timewhen the air-fuel ratio is lean than at the time when the air-fuel ratiois the stoichiometric air-fuel ratio.

In the case where the air-fuel ratio of the air-fuel mixture is lean,the thermal efficiency of the engine 1 is relatively high, which reducesthe temperature of the exhaust gas. The model 803 at the time when theair-fuel ratio is lean is configured such that the temperature of theexhaust gas at the same degree of the progress of the combustion isestimated to be lower than that in each of the models 801, 802 at thetime when the air-fuel ratio is the stoichiometric air-fuel ratio. Thus,when the air-fuel ratio is the stoichiometric air-fuel ratio, and whenthe air-fuel ratio is lean, the estimation module 106 can accuratelyestimate the temperature of the exhaust gas.

The model 803 at the time when the air-fuel ratio of the air-fuelmixture is lean is configured such that the temperature increasing rateof the exhaust gas with respect to the change in the progress of thecombustion is lower than the temperature increasing rate of the model802 at the time when the air-fuel ratio is the stoichiometric air-fuelratio. That is, the gradient of the model 803 is shallower than that ofthe model 802.

As described above, the thermal efficiency is high in the SPCCIcombustion. When the air-fuel ratio of the air-fuel mixture is lean,thermal efficiency of the engine 1 is further increased. In the SPCCIcombustion, a case where the air-fuel ratio of the air-fuel mixture isthe stoichiometric air-fuel ratio and a case where the air-fuel ratiothereof is lean are compared. In such a case, an amount of the fuelsupplied to the cylinder is small when the air-fuel ratio is lean, andthe heat generation amount in the cylinder is also small when theair-fuel ratio is lean. The heat generation amount in the cylinder issmall. Thus, the change in the temperature of the exhaust gas withrespect to the delay in the progress of the combustion is small when theair-fuel ratio of the air-fuel mixture is lean. Thus, in the model 803at the time when air-fuel ratio of the air-fuel mixture is lean, thetemperature increasing rate of the exhaust gas with respect to thechange in the progress of the combustion is lower than the temperatureincreasing rate of the model 802 at the time when the air-fuel ratio isthe stoichiometric air-fuel ratio.

The model is changed according to the speed of the engine 1. FIG. 9exemplifies models 901, 903 when the speed of the engine 1 is high andmodels 902, 904 when the speed of the engine 1 is low. An upper graph inFIG. 9 is a model in the SI combustion, and a lower graph therein is amodel in the SPCCI combustion.

As illustrated in the upper graph in FIG. 9, the model 901 of the casewhere the speed of the engine 1 is high is configured such that thetemperature of the exhaust gas with respect to the same degree of theprogress of the combustion is estimated to be higher than that in themodel 902 of the case where the speed of the engine 1 is low. In thecase where the progress of the combustion is the same, the estimationmodule 106 estimates the temperature of exhaust gas to be higher whenthe speed of the engine 1 is high than when the speed thereof is low.

The temperature of the exhaust gas is higher when the speed of theengine 1 is high than when the speed thereof is low. This is because thenumber of combustions per unit time is increased with the increase inthe speed of the engine 1, thus a cylinder wall temperature isincreased, and the cooling loss is suppressed due to a short combustiontime in the cylinder.

Each of the models 901, 902 is configured such that the temperature ofthe exhaust gas with respect to the same degree of the progress of thecombustion is estimated to be higher when the speed of the engine 1 ishigh than when the speed thereof is low. In this way, even in the casewhere the speed of the engine 1 is changed, the estimation module 106can accurately estimate the temperature of the exhaust gas.

The same applies to the SPCCI combustion as illustrated in the lowergraph in FIG. 9. That is, the model 903 of the case where the speed ofthe engine 1 is high is configured such that the temperature of theexhaust gas with respect to the same degree of the progress of thecombustion is estimated to higher than that in the model 904 of the casewhere the speed thereof is low. In this way, even in the case where thespeed of the engine 1 is changed during the SPCCI combustion, theestimation module 106 can accurately estimate the temperature of theexhaust gas.

In addition, as illustrated in the upper graph in FIG. 9, the model 901of the case where the speed of the engine 1 is high is configured suchthat the temperature increasing rate of the exhaust gas with respect tothe change in the progress of the combustion is higher than that in themodel 902 of the case where the speed of the engine 1 is low. That is,the gradient of the model 901 is steeper than that of the model 902.

As described above, in the case where the speed of the engine 1 is high,the cylinder wall temperature is high, and a difference between thein-cylinder temperature and the cylinder wall temperature is relativelysmall. In the case where the speed of the engine 1 is high and where thein-cylinder temperature is changed in conjunction with the change in theprogress of the combustion, a ratio of a variation in the in-cylindertemperature to the difference between the in-cylinder temperature andthe cylinder wall temperature during the combustion is increased. Thatis, in the case where the speed of the engine 1 is high and the progressof the combustion is changed, the cooling loss is significantly changed.Thus, the temperature of the exhaust gas is also significantly changed.On the contrary, in the case where the speed of the engine 1 is low, thecylinder wall temperature is low, and the difference between thein-cylinder temperature and the cylinder wall temperature during thecombustion is relatively large. Even in the case where the speed of theengine 1 is low and where the in-cylinder temperature is changed due tothe change in the progress of the combustion, the change in thetemperature of the exhaust gas is small due to the small ratio of thevariation in the in-cylinder temperature to the difference between thein-cylinder temperature and the cylinder wall temperature.

Each of the models 901, 902 is configured such that the temperatureincreasing rate of the exhaust gas with respect to the change in theprogress of the combustion is higher when the speed of the engine 1 ishigh than when the speed thereof is low. In this way, the estimationmodule 106 can accurately estimate the temperature of the exhaust gaseven when the speed of the engine 1 is changed.

Furthermore, as illustrated in the lower graph in FIG. 9, also in theSPCCI combustion, the model 903 of the case where the speed of theengine 1 is high is configured such that the temperature increasing rateof the exhaust gas with respect to the change in the progress of thecombustion is higher than that in the model 904 of the case where thespeed of the engine 1 is low. That is, the gradient of the model 903 issteeper than that of the model 904.

Here, FIG. 10 exemplifies a relationship between the speed of the engine1 and the estimated temperature of the exhaust gas with the certainengine load. As illustrated in FIG. 4, the speed N1 of the engine 1 inthe operation map 401 is the speed at which the combustion mode isswitched between the SPCCI combustion and the SI combustion.Accordingly, the engine 1 performs the SPCCI combustion on a left sideof N1 in FIG. 10, and the engine 1 performs the SI combustion on a rightside of N1.

In the case where the speed of the engine 1 is lower than N1, theprogress of the combustion in the SPCCI combustion is delayed with theincrease in the speed. As a result, the temperature of the exhaust gasis gradually increased.

When the speed of the engine 1 is N1, the combustion mode is switchedbetween the SPCCI combustion and the SI combustion. The progress of thecombustion in the SI combustion is advanced in comparison with theprogress of the combustion in the SPCCI combustion. Thus, thetemperature of the exhaust gas is reduced when the SPCCI combustion isswitched to the SI combustion. Meanwhile, the temperature of the exhaustgas is increased when the SI combustion is switched to the SPCCIcombustion.

In the case where the speed of the engine 1 is equal to or higher thanN1, the progress of the combustion in the SI combustion is also delayedas the speed is increased. Thus, the temperature of the exhaust gas isgradually increased. However, the temperature increasing rate of theexhaust gas with respect to the increase in the speed of the engine 1 islower in the SI combustion than in the SPCCI combustion. That is, thegradient of the line in FIG. 10 is gentle. This is because, in the SIcombustion, the amount of the delay in the progress of the combustionwith respect to the increase in the speed of the engine 1 is small.

The SPCCI combustion, which significantly depends on the temperature anda pressure state inside the cylinder, is likely to be affected by thechange in the in-cylinder volume during the power stroke. In the SPCCIcombustion, the amount of the delay in the progress of the combustionwith respect to the speed of the engine 1 is large. Thus, in the SPCCIcombustion, the temperature increasing rate of the exhaust gas withrespect to the increase in the speed is high. That is, the gradient ofthe line in FIG. 10 is steep.

FIG. 11 exemplifies models 1001, 1003 when the load of the engine 1 islarge and models 1002, 1004 when the load of the engine 1 is small. Themodel is changed according to the load of the engine 1. An upper graphin FIG. 11 includes the models in the SI combustion, and a lower graphtherein includes the models in the SPCCI combustion.

As illustrated in the upper graph in FIG. 11, the model 1001 of the casewhere the load of the engine 1 is large is configured such that thetemperature of the exhaust gas with respect to the same degree of theprogress of the combustion is estimated to be higher than in the model1002 of the case where the load is small. In the case where the progressof the combustion is the same, the estimation 4 module 106 estimates thetemperature of exhaust gas to be higher when the load of the engine 1 islarge than when the load thereof is small.

When the load of the engine 1 is large, the amount of the fuel suppliedto the cylinder is increased, and thus, the heat generation amount inthe cylinder is increased. The models 1001, 1002 are configured suchthat the temperature of the exhaust gas with respect to the same degreeof the progress of the combustion is estimated to be higher when theload of the engine 1 is large than when the load thereof is small. Thus,the estimation module 106 can accurately estimate the temperature of theexhaust gas even when the load of the engine 1 is changed.

The same applies to the SPCCI combustion as illustrated in the lowergraph in FIG. 11. More specifically, the model 1003 of the case wherethe load of the engine 1 is large is configured such that thetemperature of the exhaust gas with respect to the same degree of theprogress of the combustion is estimated to be higher than in the model1004 of the case where the load is small. In this way, in the SPCCIcombustion, even when the load of the engine 1 is changed, theestimation module 106 can accurately estimate the temperature of theexhaust gas.

In addition, as illustrated in the upper graph in FIG. 11, the model1001 of the case where the load of the engine 1 is large is configuredsuch that the temperature increasing rate of the exhaust gas withrespect to the change in the progress of the combustion is higher thanthat in the model 1002 of the case where the load of the engine 1 issmall. That is, the gradient of the model 1001 is steeper than that ofthe model 1002.

When the load of the engine 1 is large, the fuel supply amount is large.In the case where the load of the engine 1 is large and the progress ofthe combustion is changed, the variation in the illustrated work becomessignificant. As a result, the temperature of the exhaust gas is alsosignificantly changed. On the contrary, in the case where the load ofthe engine 1 is small, the fuel supply amount is small. Thus, even inthe case where the progress of the combustion is changed, the variationin the illustrated work is insignificant. As a result, the change in thetemperature of the exhaust gas is small.

The models 1001, 1002 are configured such that the temperatureincreasing rate of the exhaust gas with respect to the change in theprogress of the combustion is higher when the load of the engine 1 islarge than when the load of the engine 1 is small. In this way, even inthe case where the load of the engine 1 is changed, the estimationmodule 106 can accurately estimate the temperature of the exhaust gas.

As illustrated in the lower graph in FIG. 11, also in the SPCCIcombustion, the model 1003 of the case where the load of the engine 1 islarge is configured such that the temperature increasing rate of theexhaust gas with respect to the change in the progress of the combustionis higher than that of the model 1004 of the case where the load of theengine 1 is small. That is, the gradient of the model 1003 is steeperthan that of the model 1004.

Referring back to the flowchart in FIG. 7, in step S6, the estimationmodule 106 estimates the temperature of the exhaust gas by using themodel corresponding to the SI combustion (for example, the model 801).In step S9, the estimation module 106 estimates the temperature of theexhaust gas by using the model corresponding to the case where theair-fuel ratio of the air-fuel mixture is the stoichiometric air-fuelratio in the SPCCI combustion (for example, the model 802). In step S11,the estimation module 106 estimates the temperature of the exhaust gasby using the model corresponding to the case where the air-fuel ratio ofthe air-fuel mixture is lean in the SPCCI combustion (for example, themodel 803).

In step S7 following step S6, the correction module 107 of the ECU 10corrects the temperature of the exhaust gas, which is estimated in stepS6, according to the temperature of the engine 1. In step S10 followingstep S9, the correction module 107 corrects the temperature of theexhaust gas, which is estimated in step S9, according to the temperatureof the engine 1. In step S12 following step S11, the correction module107 corrects the temperature of the exhaust gas, which is estimated instep S11, according to the temperature of the engine 1.

The temperature of the engine 1 is represented by the temperature of thecoolant for the engine 1. The correction module 107 acquires thetemperature of the coolant for the engine 1 from the signal of thesecond coolant temperature sensor SW10.

FIG. 12 exemplifies a relationship between the temperature of thecoolant and a correction amount of the temperature of the exhaust gas.T1 represents a reference temperature of the coolant. Theabove-described models 801, 802, 803, 901, 902, 903, 904, 1001, 1002,1003, 1004 each represent the relationship between the progress of thecombustion and the temperature of the exhaust gas in the case where thecoolant is at the reference temperature T1.

When the temperature of the coolant is T1, the correction amount iszero. The correction module 107 does not actually correct thetemperature of the exhaust gas estimated in steps S6, S9, S11. When thetemperature of the coolant is higher than T1, the correction amountbecomes larger than zero. As the temperature of the coolant isincreased, the positive correction amount is increased. The correctionmodule 107 makes correction to increase the temperature of the exhaustgas estimated in steps S6, S9, S11. As the temperature of the engine 1is increased, the cooling loss is reduced. Thus, the temperature of theexhaust gas is increased. When the correction module 107 makes thecorrection to increase the temperature of the exhaust gas, the ECU 10can further accurately estimate the temperature of the exhaust gas.

When the temperature of the coolant is lower than T1, the correctionamount becomes smaller than zero. As the temperature of the coolant isreduced, the negative correction amount is increased. The correctionmodule 107 makes correction to reduce the temperature of the exhaust gasestimated in steps S6, S9, S11. As the temperature of the engine 1 isreduced, the cooling loss is increased. Thus, the temperature of theexhaust gas is reduced. When the correction module 107 makes thecorrection to reduce the temperature of the exhaust gas, the ECU 10 canfurther accurately estimate the temperature of the exhaust gas.

The correction amount is changed according to the air-fuel ratio of theair-fuel mixture. More specifically, a straight line 1202 in FIG. 12represents the correction amount of the case where the air-fuel ratio ofthe air-fuel mixture is stoichiometric air-fuel ratio in the SPCCIcombustion. A straight line 1203 represents the correction amount of thecase where the air-fuel ratio of the air-fuel mixture is lean in theSPCCI combustion. As understood from FIG. 12, in the case where theair-fuel ratio of the air-fuel mixture is the stoichiometric air-fuelratio, the correction amount of the temperature of the exhaust gas isincreased to be larger than in the case where the air-fuel ratio islean.

In the case where the air-fuel ratio of the air-fuel mixture is lean,thermal efficiency of the engine 1 is relatively high, which reduces theamount of the fuel supplied to the cylinder. As a result, thetemperature inside the cylinder during the combustion is lower when theair-fuel ratio of the air-fuel mixture is lean than when the air-fuelratio is the stoichiometric air-fuel ratio. In addition, the coolingloss is also reduced to be smaller when the air-fuel ratio is lean thanwhen the air-fuel ratio is the stoichiometric air-fuel ratio. On thecontrary, in the case where the air-fuel ratio of the air-fuel mixtureis the stoichiometric air-fuel ratio, the cooling loss is increased.

The correction module 107 makes the correction in consideration of theinfluence of the cooling loss. Thus, in the case where the air-fuelratio of the air-fuel mixture is the stoichiometric air-fuel ratio, thecorrection amount of the temperature of the exhaust gas is increased tobe larger than in the case where the air-fuel ratio is lean. In thisway, the correction module 107 can appropriately correct the temperatureof the exhaust gas in consideration of the relationship between theair-fuel ratio of the air-fuel mixture and the cooling loss.

The correction amount is changed according to the combustion mode. Morespecifically, a straight line 1201 in FIG. 12 represents the correctionamount in the SI combustion. As understood from FIG. 12, the correctionamount of the temperature of the exhaust gas is increased to be largerin the SI combustion than in the SPCCI combustion.

In the SPCCI combustion, the heat generation rate in the CI combustionis high. Thus, a peak temperature inside the cylinder is higher than thepeak temperature in the SI combustion. In the SPCCI combustion, adifference between the peak temperature and the temperature of thecoolant is significant. Thus, a ratio of a temperature change amount ofthe coolant to the difference between the peak temperature and thetemperature of the coolant in the case where the temperature of thecoolant is changed is small. Thus, in the SPCCI combustion, even whenthe temperature of the coolant is changed, the amount of the coolingloss is not significantly changed. In the SPCCI combustion, even whenthe temperature of the coolant is changed, the temperature of theexhaust gas is not significantly changed.

Meanwhile, in the SI combustion, the difference between the peaktemperature and the temperature of the coolant is small. Thus, the ratioof the temperature change amount of the coolant to the differencebetween the peak temperature and the temperature of the coolant in thecase where the temperature of the coolant is changed is large. In the SIcombustion, when the temperature of the coolant is changed, the coolingloss is significantly changed. In the SI combustion, when thetemperature of the coolant is changed, the temperature of the exhaustgas is significantly changed.

Thus, in the SI combustion, the correction amount of the temperature ofthe exhaust gas is increased to be larger than that in the SPCCIcombustion. In this way, the correction module 107 can appropriatelycorrect the temperature of the exhaust gas in consideration of therelationship between the combustion mode and the cooling loss.

Referring back to the flowchart in FIG. 7, in step S7, the correctionmodule 107 corrects the temperature of the exhaust gas according to thecorrection amount 1201 corresponding to the SI combustion. In step S10,the correction module 107 corrects the temperature of the exhaust gasaccording to the correction amount 1202 corresponding to the case wherethe air-fuel ratio of the air-fuel mixture is the stoichiometricair-fuel ratio in the SPCCI combustion. In step S12, the correctionmodule 107 corrects the temperature of the exhaust gas according to thecorrection amount 1203 corresponding to the case where the air-fuelratio of the air-fuel mixture is lean in the SPCCI combustion.

By the procedure that has been described so far, the ECU 10 that hasestimated the temperature of the exhaust gas executes the control toreduce the temperature of the exhaust gas in the case where theestimated temperature of the exhaust gas is higher than the referencetemperature. The ECU 10 increases the amount of the fuel supplied to thecylinder to be larger than that when the temperature of the exhaust gasis equal to or lower than the reference temperature, for example. Whenthe amount of the fuel supplied to the cylinder is increased, due tolatent heat of the fuel, the amount of which is increased, thetemperature of the exhaust gas discharged from the cylinder is reduced.By reducing the temperature of the exhaust gas to be lower than thereference temperature, it is possible to secure reliability of thecatalyst provided in the exhaust passage 50 of the engine 1.

The ECU 10 may reduce the temperature of the coolant supplied to theengine 1 to be lower than that in the case where the temperature of theexhaust gas is equal to or lower than the reference temperature. Morespecifically, the ECU 10 controls the thermostat valve 76 and the flowrate regulator valve 77 in the cooling system 70 so as to regulate thetemperature of the coolant supplied to the engine 1. In this way, thecooling loss of the engine 1 can be regulated, and it is possible toreduce the temperature of the exhaust gas that is discharged from thecylinder.

The ECU 10 may further increase the flow rate of the coolant supplied tothe engine 1 to be higher than that of the case where the temperature ofthe exhaust gas is equal to or lower than the reference temperature.More specifically, the ECU 10 controls the thermostat valve 76 and theflow rate regulator valve 77 in the cooling system 70 to regulate theflow rate of the coolant supplied to the engine 1. In this way, thecooling loss of the engine can be regulated, and it is possible toreduce the temperature of the exhaust gas that is discharged from thecylinder.

The above-described control for reducing the temperature of the exhaustgas can be combined.

In the configuration example, the ECU 10 uses the models, each of whichrepresents the relationship between the progress of the combustion andthe temperature of the exhaust gas, to estimate the temperature of theexhaust gas, and corrects the estimated temperature of the exhaust gasaccording to the temperature of the engine 1. Differing from the above,a model that represents a relationship among the progress of thecombustion, the temperature of the exhaust gas, and the temperature ofthe engine may be created. Then, the ECU 10 may use such a model toestimate the temperature of the exhaust gas from the progress of thecombustion and the temperature of the engine. This configurationcorresponds to a configuration in which the correction module 107 isprovided in the estimation module 106.

The ECU 10 may estimate the temperature of the exhaust gas by using amap that at least represents the relationship between the progress ofthe combustion and the temperature of the exhaust gas instead of themodel that represents the relationship between the progress of thecombustion and the temperature of the exhaust gas.

The technique disclosed herein is not limited to the technique appliedto the engine 1 having the above-described configuration. Any of variousconfigurations can be adopted as the configuration of the engine 1. Forexample, the technique disclosed herein may be applied to a dieselengine in which the in-cylinder air-fuel mixture is not forciblyignited. Also, with the diesel engine, the control unit can accuratelyestimate the temperature of the exhaust gas from the model, whichrepresents the relationship between the progress of the combustion andthe temperature of the exhaust gas, and the progress of the combustion.

It should be understood that the embodiments herein are illustrative andnot restrictive, since the scope of the invention is defined by theappended claims rather than by the description preceding them, and allchanges that fall within metes and bounds of the claims, or equivalenceof such metes and bounds thereof, are therefore intended to be embracedby the claims.

DESCRIPTION OF REFERENCE CHARACTERS

1: Engine

10: ECU (control unit)

105: Calculation module

106: Estimation module

50: Exhaust passage

801, 802, 803: Model

901, 902, 903, 904: Model

1001, 1002, 1003, 1004: Model

SW6: In-cylinder pressure sensor

SW11: Crank angle sensor

The invention claimed is:
 1. An engine controller for an engine, theengine controller comprising: an exhaust passage which is connected tothe engine and through which exhaust gas is discharged from inside of acylinder of the engine; a sensor which outputs a signal corresponding toa combustion state in the cylinder; and a control unit to which thesensor is connected, which estimates a temperature of the exhaust gas onthe basis of the signal of the sensor, and which controls the engineaccording to the estimated temperature of the exhaust gas, wherein thecontrol unit changes an air-fuel ratio of an air-fuel mixture in thecylinder to a stoichiometric air-fuel ratio or a leaner air-fuel ratiothan the stoichiometric air-fuel ratio according to an operation stateof the engine, the control unit includes a processor configured toexecute: a calculation module that calculates progress of combustion asa crank angle at the time when the combustion in the cylinder isprogressed to a particular extent on the basis of the signal of thesensor; and an estimation module that estimates the temperature of theexhaust gas on the basis of the progress of the combustion calculated bythe calculation module, the air-fuel ratio of the air-fuel mixture, anda temperature of the engine, and in a case where the air-fuel ratio ofthe air-fuel mixture is the stoichiometric air-fuel ratio, theestimation module estimates the temperature of the exhaust gas on thebasis of the progress of the combustion, the temperature of the engine,and a first relationship that is at least defined between the progressof the combustion and the temperature of the exhaust gas, and in a casewhere the air-fuel ratio of the air-fuel mixture is leaner than thestoichiometric air-fuel ratio, the estimation module estimates thetemperature of the exhaust gas on the basis of the progress of thecombustion, the temperature of the engine, and a second relationshipthat differs from the first relationship and is at least defined betweenthe progress of the combustion and the temperature of the exhaust gas.2. The engine controller according to claim 1, wherein the estimationmodule estimates the temperature of the exhaust gas from the progress ofthe combustion on the basis of a relationship including the firstrelationship or the second relationship, and corrects the estimatedtemperature of the exhaust gas according to the temperature of theengine, and the estimation module changes a correction amount of thetemperature of the exhaust gas according to the air-fuel ratio of theair-fuel mixture.
 3. The engine controller according to claim 2, whereinin the case where the air-fuel ratio of the air-fuel mixture is thestoichiometric air-fuel ratio, the estimation module increases thecorrection amount of the temperature of the exhaust gas to be largerthan in the case where the air-fuel ratio of the air-fuel mixture isleaner than the stoichiometric air-fuel ratio.
 4. The engine controlleraccording to claim 2, wherein the estimation module makes a correctionto reduce the estimated temperature of the exhaust gas as thetemperature of the engine is reduced in a case where the temperature ofthe engine is equal to or lower than a specified temperature, and makesa correction to increase the estimated temperature of the exhaust gas asthe temperature of the engine is increased in a case where thetemperature of the engine exceeds the specified temperature.
 5. Theengine controller according to claim 1, wherein the control unitswitches between a first combustion mode, in which the air-fuel mixturein the cylinder is forcibly ignited according to the operation state ofthe engine, so as to burn the air-fuel mixture by flame propagation, anda second combustion mode, in which the air-fuel mixture in the cylinderis forcibly ignited, so as to burn some of the air-fuel mixture byself-ignition, and in the second combustion mode, the control unitchanges the air-fuel ratio of the air-fuel mixture according to theoperation state of the engine.
 6. The engine controller according toclaim 1, wherein in a case where the temperature of the exhaust gas ishigher than a reference temperature, the control unit increases a fuelamount supplied to the cylinder to be larger than in a case where thetemperature of the exhaust gas is equal to or lower than the referencetemperature.
 7. The engine controller according to claim 1, wherein in acase where the temperature of the exhaust gas is higher than a referencetemperature, the control unit reduces a temperature of a coolantsupplied to the engine to be lower than in a case where the temperatureof the exhaust gas is equal to or lower than the reference temperature.8. The engine controller according to claim 1, wherein in a case wherethe temperature of the exhaust gas is higher than a referencetemperature, the control unit increases a flow rate of the coolantsupplied to the engine to be larger than a case where the temperature ofthe exhaust gas is equal to or lower than the reference temperature.